The present invention comprises improvements to the reduced friction piston rings described in Meacham WO/2009/033115, which is incorporated by reference in its entirety.
It is generally known to provide liquid lubricated piston ring seals in reciprocating pistons to reduce gas flow through the diametral clearance between the piston and the bore in which it reciprocates.
Compression rings of the type typically used in internal combustion engines are heat and wear-resistant hard materials such as metal, and may have metallic or ceramic coatings to improve their friction and wear properties. Compression rings are generally circular with a rectangular cross section and a small radial gap and are installed in annular grooves in the pistons. Prior to installation the outside diameter of the rings is slightly larger than the inside diameter of the bores. The radial gap allows the rings to be elastically expanded so that they can be installed in the piston grooves. When the pistons are installed in the bores, the rings are elastically compressed to the smaller bore diameter such that the radial gaps are nearly closed. The bore, the cylindrical piston sides, the rings, and the grooves are coated with a thin layer of liquid lubricant, e.g. mineral oil. In the installed condition the rings exert a moderate radial elastic force against the bore surfaces to provide a baseline sealing force. When pressure is applied to a piston and ring assembly, the pressure difference presses the rings against the low-pressure sides of the grooves, and pressurized gas flows into the ring grooves between the piston and the rings. This pressurized gas exerts outward radial force on the rings that augments the elastic baseline force.
The value of the pressure-driven outward radial force is directly proportional to the differential pressure across the ring. The inward balancing force depends on the details of the contact interface acting on a portion of the contact interface areas between the ring and the bore surface. The unbalanced portion of the force is carried as a sliding bearing contact load between the ring and the cylinder bore. At high piston speeds typical of most of the stroke, the ring-bore sliding bearing is supported hydrodynamically on a liquid lubricant film without metal-to-metal contact. Boundary lubrication with metal-to-metal contact occurs as the piston slows and reverses at the stroke ends. This results in much higher friction, and causes most of the ring wear. The simplest case is uniform contact across the entire axial ring thickness. The contact interface area pressure near the edge exposed to the high pressure gas is equal to the high pressure, and the contact interface area pressure near the other edge exposed to the low pressure gas is equal to the low pressure. The pressure at any point between the edges is intermediate between these values. The total inward radial pressure force is less than the outward radial pressure force, and the resultant ring sealing force therefore increases with increasing differential pressure. If the pressure variation between the edges is linear, the inward radial pressure force balances about half the outward radial pressure force. Small changes in variables such as ring twist can make substantial changes in the pressure variation between the edges. Twist that opens a gap towards the high pressure edge increases the total inward radial pressure force. This reduces friction, but may increase blow-by. Twist that opens a gap towards the low pressure edge decreases the total inward radial pressure force. This increases friction and may reduce blow-by. The net outward force on the ring is the vector sum of the outward radial pressure force, the inward radial pressure force, the outward radial ring elastic force and the radial friction force between the sides of the ring and the piston ring groove. This net outward force and the friction coefficient determine the friction between the ring and the bore. If also determines the load supported by the sliding bearing formed by the contact zone between the ring outside diameter and the cylinder bore. The ring friction and wear are critically dependent on the presence of liquid lubricant, the contact zone area, the piston velocity and the net outward force. Conventional piston rings represent a difficult compromise between design parameters to achieve the best possible friction and wear performance for a given application, and a principal feature of this invention is a reduced need to compromise.
Since friction increases with increasing total outward ring radial force, one piston ring design objective is to reduce this force as much as possible consistent with keeping the ring loaded against the bore to maintain a gas seal. The most direct way to reduce the outward ring sealing force of conventional rings is to reduce the ring thickness in the axial direction. This limits the maximum outward pressure force regardless of magnitude of inward balancing force. There are limits, however, on how thin the rings can be made and survive in the engine environment. Further, the top compression ring forms a significant thermal conduction path between the piston and the cooler cylinder bore that is important in keeping the piston cool, and narrow rings reduce the thermal conduction. Consequently, many engine designs have relatively thick robust rings configured such that inward radial pressure force balances a large part of the outward radial pressure force. A variety of ring cross section contours and twist conditions have been proposed to do this. A convex ring barrel bore contact surface with line contact defines the interface areas exposed to high pressure and low pressure precisely. Alternatively a tapered ring outer bore contact surface defines a substantial outer ring area exposed to high pressure during the compression and power strokes. The effective size and position of contact areas defined by convex barrel shapes or tapers are, however, affected by relatively small amounts of ring wear. One solution is application of a hard coating to limit ring wear so that the geometry is maintained during ring break-in and service. Another is to use a ring cross section in which the high pressure interface area is largely defined by a raised flange on the outer bore contact surface, resulting in a design with controlled radial force that is insensitive to ring wear.
The liquid lubricant film has an important effect on piston ring performance. During mid-stroke the high piston velocity causes the liquid lubricant to form a hydrodynamic film between the ring outer bore contact surface and the cylinder bore that prevents metal to metal contact. In this hydrodynamic mode the ring acts as a linear slider bearing of the type described in the Standard Handbook for Mechanical Engineers, 7th Edition, edited by Theodore Baumeister. Page 8-171 (1960) published by the McGraw-Hill Book Company, New York. This lubrication theory indicates that the hydrodynamic coefficient of friction fh is:
  f      h    ⁢                  ∝                  μγC        W            where μ is lubricant viscosity, γ is piston velocity, C is the bore circumference and W is the radial load supported. Fh is generally low. The coefficient of friction fh in boundary lubrication increases dramatically because of metal to metal contact between the ring and the bore. Significantly the hydrodynamic coefficient of friction is independent of axial ring bearing contact width. This theory also indicates that the lubricant film thickness ho is:
  h      o    ∝          L      ⁢                        μγC          W                    where L is the axial ring bearing contact width resulting in increased film thickness with increased contact width. The conclusion is that increased axial ring bearing contact width increases the lubricant film thickness at a given piston velocity and lubricant viscosity without increasing the friction coefficient.
As the piston slows and reverses at the stroke ends, the lubricant film thickness ho decreases, and hydrodynamic lubrication transitions to boundary lubrication with metal to metal contact and increased friction and wear. Increased bearing contact width increases lubrication film thickness ho and delays the transition to boundary lubrication. Alternatively, increased bearing contact width might be used advantageously to reduce lubricant viscosity rather than delaying the transition to boundary lubrication, allowing reduced friction elsewhere in the engine that more than offsets the increased boundary layer friction of the piston rings.
The transition to boundary lubrication may also be delayed by squeeze film lubrication. Squeeze film lubrication is a transient process in which the lubricant film is squeezed out of the gap between the ring and the cylinder bore as the surfaces approach metal-to-metal contact. The film squeezing process requires a period of time that increases with the initial gap, lubricant viscosity and contact width. This time period extends the effective hydrodynamic lubrication regime and reduces the boundary lubrication regime.
In theory increased contact width maximizes the hydrodynamic bearing regime and minimizes the boundary lubrication regime without changing the friction coefficient at constant lubricant viscosity and radial ring force. The question is how to utilize these effects. Conventional rings with increased contact width, as discussed earlier, tend to have high outward radial pressure force and high friction that negates improved bearing performance. Conventional rings therefore typically compromise towards low contact width to minimize radial force and friction through most of the stroke and accept a certain amount of boundary lubrication sliding and wear at the stroke ends. The objective of this invention is to provide piston ring configurations that effectively increase the bearing area while minimizing the radial ring force and friction, and enhance the squeeze film effect to reduce friction and wear at the stroke ends.